1. Field of the Invention
The invention generally relates to a seal for use within an intershaft assembly. Specifically, the invention is a sealing system including divergent flow grooves which separate fluid, originating from a high pressure region, communicated onto each apex along the divergent flow grooves so as to produce a balanced pressure profile radially widthwise across a piston ring disposed between concentric, rotatable inner and outer shafts. The divergent flow grooves minimize twisting along a piston ring otherwise produced by conventional hydrodynamic grooves along a mating ring. The invention is applicable to a variety of uses wherein concentric shafts are disposed in a co-rotating or counter-rotating arrangement, one specific non-limiting example being a turbine engine.
2. Background
Intershaft seal systems and hydrodynamic grooves are both known within the seal art.
Lipschitz describes a circumferential inter-seal for sealing between relatively rotatable concentric shafts in U.S. Pat. No. 4,972,986. With reference to FIG. 1, the intershaft seal assembly 1 includes a forward mating ring 4 adjacent to a high pressure region and an aft mating ring 5 adjacent to a low pressure region. Mating rings 4, 5 are disposed about a seal ring 6. A spacer ring 8 is concentrically aligned with the seal ring 6 and separated therefrom via an annular space 10. The spacer ring 8 is wider than the seal ring 6 so as to contact both forward and aft mating rings 4, 5, thus allowing the seal ring 6 to translate between the mating rings 4, 5. The mating rings 4, 5, spacer ring 8, and seal ring 6 are disposed between an inner shaft 3 and an outer shaft 2 which are concentric and rotatable in either a co-rotating or counter-rotating fashion. The mating rings 4, 5 and spacer ring 8 are secured to the inner shaft 3 via a stop ring 7 so as to rotate with the inner shaft 3. The seal ring 6 is dimensioned so as to at most partially contact the outer shaft 2 when the outer and inner shafts 2, 3 are at rest. The seal ring 6 further includes at least one circumferential space or gap which allows the seal ring 6 to flex or expand as it rotates so as to then contact and rotate with the outer shaft 2 separate from the inner shaft 3. Contact between the rotating seal ring 6 and mating rings 4, 5 is avoided to minimize friction-induced wear along the sides of the seal ring 6.
Contact between a rotating seal ring 6 and mating rings 4, 5 is minimized by a thin-film interposed between the seal ring 6 and forward mating ring 4 and between the seal ring 6 and aft mating ring 5. The thin film is produced by communicating fluid, examples including but not limited to air and air/oil mixture, from the high pressure region to the low pressure region along a path defined by spaces between the seal ring 6 and the outer and inner shafts 2, 3 and mating rings 4, 5, and axially through passages.
With reference to FIGS. 1-3, fluid flows from the higher pressure region through a plurality of horizontal ports 9 through the forward mating ring 4 and across a forward annular space 12 between the forward mating ring 4 and outer shaft 2. The same fluid flows around the seal ring 6 and through the seal ring 6 via horizontal ports 14. Thereafter, the fluid flows to the lower pressure region via an aft annular space 13 between the outer shaft 2 and the aft mating ring 5.
The forward and aft mating rings 4, 5 include a plurality of spiral grooves 11, as shown in FIGS. 2 and 3. The spiral grooves 11 include a plurality of recessed arcuate slots along the surface of the forward and aft mating rings 4, 5. The spiral grooves 11 communicate a hydrodynamic lift force onto the seal ring 6 via a pressure field along the spaces between the mating rings 4, 5 and seal ring 6. This hydrodynamic lift force increases exponentially as the seal ring 6 translates toward one of the mating rings 4, 5, thereby preventing the seal ring 6 from contacting the mating rings 4, 5 under dynamic conditions. FIGS. 4 and 5 graphically present non-symmetric pressure profiles produced by the forward and aft mating rings 4, 5, respectively, across the width of the seal ring 6.
It is well known that large non-symmetric counter forces generated by hydrodynamic grooves cause a sealing ring to twist thus compromising the parallelism between the mating rings or runners and the sealing ring. Often, the result is radial and angular distortions which produce “coning” along the sealing ring. Coning is understood to cause excessive wear along a sealing ring and to degrade the performance of a sealing system.
Lipschitz explicitly recognizes this problem and suggests for the sealing ring to be composed of materials having a high modulus of elasticity to minimize undesirable radial and angular deflections imposed by unbalanced hydrodynamic forces. As such, Lipschitz teaches away from pressure-based solutions to the twisting problem.
With reference to FIGS. 6 and 7a-7c, another intershaft seal assembly 30 is shown including a forward mating ring 33 and an aft mating ring 34 disposed about a piston ring 35 between rotatable inner and outer shafts 31, 32. The piston ring 35 is concentrically aligned with a spacer ring 36, so that the piston ring 35 translates between the forward and aft mating rings 33, 34. The piston 35 and spacer ring 36 are separated by an annular gap 40. The piston ring 35 is dimensioned and includes one or more gaps so as to flex or expand as the inner and outer shafts 31, 32 rotate, thereby contacting the outer shaft 32 and rotating therewith. The forward and aft mating rings 33, 34 and spacer ring 36 are secured to a carrier 37 via a stop ring 54 so as to rotate with the inner shaft 31. An annular space 38, 39 is provided along the forward mating ring 33 and aft mating ring 34, respectively, so as to avoid contact with the outer shaft 32. Fluid from the high pressure region passes over the forward mating ring 33, around the piston ring 35, and then over the aft mating ring 34 into the low pressure region.
In this design, conventional hydrodynamic grooves are positioned along the faces 42, 41 of the forward and aft mating rings 33, 34, respectively, to improve flow around the piston ring 35 and to minimize contact between the piston ring 35 and mating rings 33, 34. The face 41 along the aft mating ring 34 includes a plurality of outward flow hydrodynamic grooves 43, as represented in FIG. 7b. The face 42 of the forward mating ring 33 includes a plurality of inward flow hydrodynamic grooves 44, as represented in FIG. 7c. Hydrodynamic grooves 43, 44 are generally arcuate-shaped, shallow slots along the surface of the respective mating rings 33, 34.
With reference to FIG. 7d, the inward and outward flow hydrodynamic grooves 44, 43 each communicate a generally non-symmetric, triangular-shaped pressure profile 15, 16 onto opposing sides of the piston ring 35. This means that the piston ring 35 experiences higher pressures and larger unbalanced forces near the inner diameter along the forward mating ring 33 and near the outer diameter along the aft mating ring 34. The resultant deflections cause the piston ring 35 to twist 17 allowing the piston ring 35 to rub against the aft mating ring 34 so as to impart a wear pattern 18 with pronounced wear depth 45 adjacent to the inner diameter of the piston ring 35. Similar wear is likewise possible along the piston ring 35 near the outer diameter adjacent to the forward mating ring 33. In this example, material properties alone are not sufficient to avoid the distortions and wear at the higher relative rotational speeds required to further improve the performance of turbine engines with concentrically rotating shafts.
Lindeboom describes a straight leakoff seal for use within a centrifugal pump in U.S. Pat. No. 3,751,045. With reference to FIGS. 8 and 9, the centrifugal pump 20 includes a housing 21 with a collar 23 disposed about a drive shaft 22. An annular dam 26 is disposed along a sealing ring 24 toward the outer diameter thereof adjacent to the high-pressure end of the centrifugal pump 20. A plurality of v-shaped grooves 25 are equally spaced circumferentially about sealing ring 24. The v-shaped grooves 25 pump fluid from an outer region to an inner region by allowing the fluid to enter one end and exit the other end of the v-shaped grooves 25 causing the face pressure profile to grow, thus preventing contact between the sealing ring 24 and collar 23. As such, Lindeboom does not allow for apex-centric flow patterns along the v-shaped grooves 25.
Lindeboom describes the advantages of his invention via reference to the pressure profiles reproduced in FIG. 10, wherein reference “B” describes the non-hydrodynamic pressure forces acting along the back end of the sealing ring 24, reference “C” describes the hydrodynamic pressure forces acting along the interface between the collar 23 and sealing ring 24, and reference “D” describes the maximum restoring forces which prevent contact between the sealing ring 24 and collar 23 under dynamic running conditions. The pressure profiles reported by Lindeboom along the interface across the dam 26 and between the collar 23 and sealing ring 24 across the sealing ring 24 are non-symmetric. As such, Lindeboom neither suggests nor teaches the generation of a symmetric pressure profile along the width of a seal ring via symmetrically-shaped grooves.
As is readily apparent from the discussions above, the related arts do not include an intershaft seal system which minimizes twist along a seal or piston ring via the communication of a substantially symmetric pressure field across the width of the ring via a plurality of substantially symmetric hydrodynamic pockets.
Accordingly, what is required is an intershaft seal system which communicates a substantially symmetric pressure field across the width of a piston ring onto at least one side thereof via a plurality of substantially symmetric hydrodynamic pockets which receive fluid from a high pressure region and separate the flow in a divergent fashion prior to communicating the fluid onto the ring.